Regenerative heat exchanger for gas turbines



May 1958 E. T. LINDEROTH 2,836,398

REIGENERATIVE HEAT EXCHANGER FOR c'As TURBINES l2 Sheets-Sheet 1 Filed Sept. 10, 1953 TURBINE COMBUSTION CHAMBER HEAT FIG.

A/R COOLER IN VEN TOR.

ERIK TORVALD LINDEROTH mmDwmmma VOLUME ATTORNEY FIG. 4

May 27, 1958 E. T. LINDEROTH REGENERATIVE HEAT EXCHANGER FOR GAS TURBINES l2 Sheets-Sht 2 Filed Sept. 10, 1953 INVENTOR. ERIK TORVALD LINDEROTH ATTORNEY y 1953 E. T. LINDEROTH 2,836,398

REGENERATIVE HEAT EXCHANGER FOR GAS TURBINES 12 Sheets-Sheet 3 Filed Sept. 10, 1953 FIG. 9

FIG. 8.

FIG.

m M I q l w mm +&

Cold Gus L ow Pressure C d INVENTOR.

ERIK TORVALD LINDEROTH High Pressure ATTORNEY May 27, 1958 E. T. LINDEROTH REGENERATIVE HEAT EXCHANGER FOR GAS TURBINES Filed Sept. 10. 1953 12 Sheets-Sheet 4 INVENTOR. ERIK TORVALD LlNDEROTH ATTORNEY May 27, 1958 REGENERATIVE HEAT EXCHANGER FOR G'AS TURBINES Filed Sept. 10, 1953 Coefficient of Heat Transfer E. T. LINDEROTH 2,836,398

12 Sheets-Sheet 5 Q c H G. \8

- ,d 27 i Turbulent F/ow INVENTOR! ERIK TORVALD LINDEROTH Diameter of Tube FIG. l9

fluzi a AT TO RNEY May 27, '1958 E. T. LINDEROTH REGENERATIVE HEAT EXCHANGER FOR GAS TURBINES Filed Sept. 10, 1953 12 Sheets-Sheet 6 FIG. 20

mm m F m m m w.

ERIK TORVALD LINDEROTH ATTORNEY REGENERATIVE HEAT EXCHANGER FOR GAS TURBINES Filed Sept. 10, 1955 May 27, 1958 E. T. LINDEROTH 12 Sheets-Sheet 7 IN V EN TOR.

ERIK TORVALD LINDEROTH llllllllllllllllllflllllllllllifllfl FIG. 25

ATTORNEY y 1958 E. T. LINDEROTH 2,836,398

REGENERATIVE HEAT EXCHANGER FOR GAS TURBINES Filed Sept. 10, 1953 12 Sheets-Sheet 8 9 35 M mm FIG. 27

INVENTORJ RIK TORVALD LINDEROTH ATTnR NFY May 27, 1958 E. T. LINDEROTH REGENERATIVE HEAT EXCHANGER FOR GAS TURBINES Filed Sept. 10, 1953 12 Sheets-Sheet 9 FIG. 28

IN V EN TOR.

BYERIK TORVALD LINDEROTH FIG. 30

ATTORNEY y 1953 E. T. LINDEROTH 2,836,398

REGENERATIVE HEAT EXCHANGER FOR GAS TURBINES Filed Sept. 10, 1953 12 Sheets-Sheet 10 FIG. 33 INVENTOR.

ERIK TORVALD LINDEROT H ATTORNEY May 27, 1958 E. T. LINDEROTH 2,836,398

REGENERATIVE HEAT EXCHANGER FOR GAS TURBINES Filed Sept. 10, 1953 12 Sheets-Sheet 11 IN VEN TOR.

ERiK TORVALD LINDEROTH ATTORN EY y 1953 E. T. LlNDEROTH 2,836,398

REGENERATIVE HEAT EXCHANGER FOR GAS TURBINES Filed Sept. 10, 1953 12 Sheets-Sheet 12 Efficiency FlG. 35

INVEN TOR. ERIK TORVALD LINDEROTH ALTTORNEY per horse power of the output of the turbine.

United States Patent REGENERATIVE HEAT EXCHAN GER FOR GAS TURBINES Erik Torvald Linderoth, Stockholm, Sweden Application September 10, 1953, Serial No. 379,362

12 Claims. (Cl. 257-1) The present invention relates to regenerative heat exchangers for transferring heat between two gaseous media, particularly to heat exchangers for gas turbines.

Gas turbines may be operated in an open system or in a closed system. With either system, the heat exchanger serves to transfer the heat of the exhaust gases of the turbine to the air fed to the turbine for the purpose of preheating the air.

The principal object of the present invention is to provide a novel and improved heat exchanger design which has a coefficient of eificiency much higher than is attainable with heat exchangers as hitherto known. Calculations and tests have shown that a heat exchanger according to the invention is capable of attaining an efficiency of 90 percent. A gas turbine associated with a heat exchanger of such a high efliciency will obtain a fuel economy appreciably superior to that of other heat powered engines including diesel engines.

Another object of the invention is to provide a novel and improved heat exchanger design which permits to attain the aforesaid high coefficient of efficiency with a heat exchanger thatis compartively small and inexpensive. With heat exchangers and gas turbines as hitherto known and which are similar in design to those employed in connection with steam power plants, the aforesaid high efficiency is very difficult to obtain and, if obtained at all only at the expense of extreme bulk and prohibitively high costs by reason of the large volume of gas passing through a gas turbine and the large quantities of heat to be transferred from the exhaust'gases to the air at rather small differences in temperature. Calculations and tests have shown that the thermo-technical aspects of the aforesaid problem can be solved more easily by applying the principle of regenerative heat transfer, that is, by means of a mass alternatingly absorbing and yielding heat, which mass is provided with canals alternately traversed by two mediate, one yielding heat and the other absorbing heat. Earlier attempts to construct regenerative heat exchangers for gas turbines have not been successful due to the fact that too large a part of the feeding air was lost in the heat exchanger owing to sluicing and leakage.

Thus, a further object of the invention is to limitate the losses due to sluicing and leakage to permissible values. The attainment of this object presents considerable diiiiculties with heat exchangers for gas turbines by reason of the fact that the pressure difference between the gas and air sides of the heat exchanger is generally in the order of several atmospheres in contrast to steam installations in which the pressure difference is usually a few hundred millimeters of water column. In addition, every percent of air leaking away constitutes a loss of power which is several times greater with a gas turbine than with a steam installation, partly due to the higher pressure and partly due to the larger quantity of air Hence, the means of limiting the losses owing to sluicing and leakage as more fully described below are most im- 'ice portant. The possibilities of attaining the aforesaid object are intimately connected with the possibility of obtaining the required heat transfer with an extremely small mass provided with narrow gauge flow canals.

Other and further objects, features and advantages of the invention will be pointed out hereinafter and set forth in the appended claims forming part of the application.

In the accompanying drawings several now preferred embodiments of the invention are shown by way of illustration and not by way of limitation.

In the drawings:

Fig. 1 is a block diagram of an installation for a gas turbine operated according to the open system and in which the air preheated by a heat exchanger according to the invention is fed to the combustion chamber.

Fig. 2 is a block diagram of an installation for a gas turbine in which the turbine is operated according to the closed system and in which the air preheated by the heat exchanger according to the invention is fed to a further air heater heated by an external source of heat before reaching the turbine.

Pig. 3 is a block diagram of an open system similar to Fig. 1 out in which the air preheated by the heat exchanger is fed directly to the turbine, the external source of heat being located after the turbine.

Pig. 4 is a pressure-volume diagram of a turbine.

Fig. 5 is a sectional top view of a heat exchanger according to the invention.

Fig. 6 is a section taken on line 66 of Fig. 5.

Figs. 7 and 8 are fragmentary sectional views of the heat transfer surface of a heat exchanger according to Figs. 5 and 6, Fig. 7 showing the transfer surface in natural size and Fig. 8 is an enlargement of said surface.

Figs. 9 and 10 are fragmentary sectional views of a modification of the heat transfer surface for the heat exchanger according to Figs. 5 and 6, also showing the heat surface in natural size and on an enlarged scale respectively.

Figs. 11 and 12 are fragmentary sectional views of another modification of the heat transfer surface for the heat exchanger according to Figs. 5 and 6, again showing the said surface in natural size and on an enlarged scale respectively.

Fig. 13 is a fragmentary view of sealing and soot blowing means, partly in section.

Fig. 14 is a section taken on line 14-14 of Fig. 13.

Fig. 15 is a modification of the sealing means of Fig. 13 in the same section.

Fig. 16 is a sectional fragmentary view of sealing and guiding means of the heat exchanger.

Fig. 17 is a section taken on line 17-17 of Fig. 16.

Fig. 18 is a fragmentary sectional view of sealing and cooling means of the heat exchanger according to the invention.

Fig. 19 is a graph of the heat transfer factor a as a function of the diameter of a tube through which hot gas is flowing.

Fig. 20 is a fragmentary plan view, partly in section, of a modification of a heat exchanger according to the invention.

Fig. 21 is a section taken on line 21-21 of Fig. 20.

Figs. 22 and 23 are views similar to Fig. 21 but showing modifications of the arrangement according to this figure.

Fig. 24 is a sectional view of a device for removing precipitations from the heat transfer surface according to the invention.

Fig. 25 is an enlarged detail view of Fig. 24.

Fig. 26 is a view, partly in section, taken on line 2626 of Fig. 25. a

:being. fed to'the combustion chamber; operating with low pressure and high temperature the.

Figs. 28 to 31 are diagrammatic views illustrating the phenomena. utilized .for :the removal of precipitations V 'withthe device according toFigs';v 24 and 27. V

f. Fig. 32 isan elevational sectional view of still another modification of aheat exchanger according to the invention. H

j Fig. 33 is asection taken on line 3333 of .Fig. 32.

Fig. 34 isa fragmentary view of a detail view of Fig.32

' on an enlarged scale, and

..Fig. 35 is agraph showing thetheoretical maximum eflic'iencyof a'heatexchanger according to the invention provided with a heat .transfer'surface according to Fig. 34-.

Referring in detail to Figs. 1, 2 and 3,' thelgasturbine 'is operated according to the so-called constant-pressure principle. .In the open system according to Fig. 1 the turbine ldrives acompressor 2 which compresses the and serves to transfer the'heat of the exhaust gases from the turbine to the air so that: the air is preheated before In gas turbines ;major portion ofthe heat requiredfor preheating the air is recovered from theexhaust gases sothat only a small portion .of the required" heat must be supplied to the sys- -ten11from the outside by means of fuel as .will be more fully explained hereinaftenl N Fig. 27 isa section taken on line 27-27 of Fig. 24 f on an enlarged scale.

In the closed system according to Fig. 2, the air circu lates within the system in a closed circuit and is again preheated'in heat exchanger 4. However, before being .fed' to. the turbine 1'it is further heated in an air heater .5 which inay'be a recuperative heater in which the air flows along one side of a heated plate or tube, Beforethe air'fromthe turbine is returned to compressor 2 it must be cooled in an air 'co,oler 6. exchanger 4 in such a closed system hasthe advantage that the amount of heat which must be delivered by' the air heater 5 is reduced by the recovery of heat from the exhaust gases, of the turbine and the amount of heat which must; be removed from the :systemby the cooler 6 will also be reduced. As a result, the required heat transfer surfaces in the heat exchangers 5 and 6 are reduced correspondingly, that is,- an ,;expensive and inefiective heating surface is replaced by aninexpensive and effective one. 7 At the same time fuelis saved bytherecovery. of heat fromthe airj exhausted from theturbinej, o

In the open system according toFi'g: 3- the combustion chamber 3 is p1aced behind the turbine, Consequently, no combustion takes place; ahead of the turbine and the latter'is fedwith air preheated in heatexchanger 4 to the temperature required for the operation of the turbine. As aresult of the expansion ofthe heated air within the turbine, the temperature of the air decreases, heat being converted to'work, and' thecorresponding amount of heat isadded in the combustion chamber 3'which heat is pro duced by fuel. The heat generated in the combustion chamber is transferred in-theheat exchanger to the'air flowing through the same and fed to the turbine; 1 The advantage of this system is that the turbine operates With air asit does in. the closed system of Fig. 2 andthat the cooler 6 of the system of Fig. 2 is not necessary. How-' ever, very high demands are'made upon the :heat ex changer. i

Also the heat exchangersof Figs. 1 and 2which heat from the exhaust side of the turbine must satisfy very high requirements as to the transfer ofheat'; The vol ume of gas leaving a gas' turbine is about 3 to 5 times higher than the volume of gas from the exhaustside of a steam power'plant for the corresponding magnitude of the delivered power. Furthermore, the temperature The useof the heat As the result of the elevation of the air temperature, I the volume of the air increases in direct proportion to oflthe gas coming, from a gas turbine isconsiderablyf higherthan the temperature of the exhaust gases from a f tional heat from the exhaustgases' of the turbine is cor respondingly reduced. Consequently, it is desirable for 'a gocdfuel economy that the heat exchanger is capable of cooling the exhaust gases from'the turbine to a'temperature slightly higher-than the -teniper ature of the air coming from the "compressor which air in turn should be heated to a temperature close to the temperature of the exhaust gases from the turbine. One 7 of the requirements to satisfy the aforementioned conditions is to conduct the gas and the air acting upon'each otherin countercurrent.

The difiiculties present in the design of a heat exchanger 7 capable of producing the aforementioned efliciency of under the above outlined conditions will befurther explained in connection with the diagram of Fig. 4.

This figure shows a co-called pressure-volume diagram f I for a gas turbine; operating according to vtheconstantpressure system and applies in principle to systems ac- I cording to Figs. 1, 2 and 3.- A relatively low' compression has been selectedto show that a high efiiciency can i V be attainedby means of the heat exchanger according .to the invention'in spite of the low compression.

' ,Let it be first assumed that the turbine is operated in an open system and that the compressor drawsin air from the atmosphere at. atemperature of 20 C., point A of the diagram, and that the air iscompressed to two atmospheres above atmospheric pressure'so that the total air pressure is threeatmospheres. I is heated in the compressor to a temperature'of about 140 C., at which temperature the air is delivered to the combustion chamber; point'B. In the combustion chamber the temperature of the air is elevated by thecombustion of fuel to a temperature of 800 C., ,point'C.

the increase in temperature (constant-pressure), calculated in absolute temperature, so that the volume of the air at pqintC is 2.6j timesthe volume of the air atopoint B. The gas expands in. thelturbine to atmospheric pressure whereby work'is performed in the'turbine, the temperature decreasing to 5 5 0 C., point D.

'The factorsto be satisfied in the design of a heat exchanger for thef'sy'stems of Figs. 1 and 2'(recovery of heat from exhaust. gas). which is eflicient aceording .to the diagram of Figs 4 shall be first analyzed. As stated before, the heat" exchanger should be capable of heating air having an initial temperature of .140" C. by gases having a temperature of 550 C. Hence, the: theoretical maximum temperature, of the preheated 'air is 550 C., assuming an infinitely high heat transferfactor and an ideal or full 'countercurrent, and the theoretical lowest temperature to which the gascan'be cooledis 140 C.

assuming a percent efliciency of the heat exchanger. On the basis ofthese figures, a heatexchanger with the desired 90 percent efificiency produce. the following results: 7 a w 7 ss heated from @510 c. Gas is'cooledfrorn 550 to 1 80 C.

Consequently, the total efliciency of the power generation is raised from 20 percent to'45 percent, assuming a 100 percent" combustion ofthe fuel, in the system according toFig. 1."

h 'A' gas turbine prQducingQiOOO and having .an efiiciency of both the compressor and the turbine of 87 percent consumes about 30 kg. ofair per-second at' the This air isheated to a certain extent by s. a resultlth a r pressure and the temperatures according to the diagram of Fig. 4. This means that a heat exchanger operating at the aforementioned ranges of temperature must trans mit million kcal. per hour, and this great quantity of heat must be transferred at a difference of temperature of only 40 C. between the gas and the air.

The requirements are still higher for a heat exchanger to be used in the system of Fig. 3 for a turbine also producing 5,000 H. P. The pertinent data are as follows:

30 kg. of air per second must be heated from 140 to 30 kg. of gas must be cooled per second from 840 to The difierence of temperature is again assumed to be 40 C. so that the efiiciency of the entire system will remain the same. However, the efliciency of the heat exchanger must be increased to 94 percent, and the quantity of heat to be transferred is increased to 18 million kcal. per hour.

It will be clear from these examples that new ways of solving this problem have to be found.

According to my prior U. S. Patents Nos. 2,227,836 and 2,178,481 a type of heating surface is suggested for regenerative heat exchangers in which the flow of gas is split up in very thin layers in a. heat transfer surface provided with very narrow and short canals for instance, fine meshed wire gauze or perforated plates.

At the time when those patents were filed, gas turbines were still in their experimental stage and not until later years were gas turbines commercially employed. On the other hand, the designs of the aforementioned patents mentioned could not be applied to steam power plants as the opinion was held that the high degree of impurity of the flue gases prevented the use of the basic features of I the patents. For gas turbine however-owing to the great sensitivity of the turbine to impurity in the driving gasit was necessary to use either fuels not producing appreciable impurities or thoroughly to purify the gas before feeding it to the turbine.

One of the difficulties confronting the designer of regenerative air preheaters in connection with gas turbines resides in keeping air and gas apart without too great an intermixture. The aforesaid patents disclose regenerative heat exchangers in which the heat transfer surface rotates in a housing divided into two chambers, one for flue gas and one for air, every part of the heat transfer surface alternatingly passing the gas chamber and the air chamber respectively. The stationary partition wall between the gas side and the air side must not be in contact with the rotor but a certain clearance should be left sufliciently large to compensate for the deformations of the rotor due to heat expansion. For steam power plants the said clearance is not equally important, as the pressure differences between the gas side and the air side in the heat exchanger are small compared'with those of the gas turbine. With steam power plants, the pressure is calculated in millimeters of water column and with gas turbines in atmospheres.

In spite of the small pressure differences with steam power plants, the air leakage of the regenerative heat exchangers for this purpose is of the magnitude of 6 to percent. A leakage of such magnitude cannot be tolerated with gas turbines as it would imply too great a loss of power. A maximum leakage of 2 percent could be tolerated if the heat exchanger has such a high degree of efliciency as the one exemplified above, and the turbine works with high degrees of temperature.

The conditions for satisfying such high requirements are unfavorable not only due to the great differences of pressure but also due to the high temperature and the demands for a high degree of heat efliciency. The high temperature will cause the rotor to warp owing to the temperature tensions, and the high degree of heat efii- 6 ciency requires large heat surfaces entailing a large rotor and great lengths of the gaps to be sealed as well as a large volume of the rotor which in turn involves that large quantities of air will be lost by sluicing during the rotation of the heat surface.

An effective heat surface permitting a radical decrease of the dimensions of the rotor is consequently not only a problem dependent on the sealing but it is the principal requirement for reducing the losses caused by leakage and sluicing.

Calculations have shown that by merely reducing the heating surface, leakage cannot be reduced to the degree required. With the most common type of heat exchangers of the regenerative type-Ljungstroms air preheaterthe heat transfer surface is supported by a rotor frame shaped like a wheel with plate like spokes facing with their edges the side surface of the wheel. Between these spokes the heat transfer surface is placed in chambers. The flow directions of air and gas are axial and opposed to each other. The rotor is carried at its center and slowly rotated in a housing divided into two chambers, one for gas and one for air. Due to the countercurrent, one side of the rotor frame will always be in contact with the hot gas and the preheated air while the opposite side of the rotor will always be in contact with the cold air and the cooled gas. Owing to this fact, heat tensions will arise which will deform the rotor.

The higher the degree of efliciency according to which the heat exchanger is to be dimensioned, the greater will be the diflerence in temperature between the two sides of the rotor. Thorough calculations of the deformations have shown the extreme difliculty if not impossibility of sealing a heat exchanger rotor designed in the usual way under the working conditions prevailing in a gas turbine plant.

According to the present invention a new principle of design is used involving quite another and more advantageous starting point for the solution of the sealing problem.

The basic principles employed according to the present invention may be outlined as follows:

The conventional rigid rotor of the heat exchanger is replaced by a flexible rotor supported on rollers. The flexibility of the rotor is such that the rotor will always rest upon the rollers, either by its own weight or due to deformation caused by pairs ofsealing rollers on opposite sides of the rotor. Consequently, tight sealing means can be provided adjacent to the rollers and to the rotor as the position of the rotor will not be altered in the proximity of the rollers by the effect of the heat. The rolls may in turn be made to seal also against the stationary parts (the partition wall). This is not a diflicult problem, since the rolls have small sizes compared with the rotor. As they are not exposed to considerable variations in temperature, they can be mounted in housmgs providing a very small clearance only and they can be further sealed by means of labyrinth sealings. V

. The rotor is preferably in form of a flexible ring rotating between the tight sealing means and so designed that the width of the heat transfer surface of the rotor 1n the direction of the flow of the gas or the air is less than 5 percent of the outer diameter of the rotor, preferably between 2.5 and 0.25 percent of the diameter. As a result, the length in the said direction of the surface to be sealed against the stationary parts of the heat exchangers is short which facilitates the sealing of the rotor. On the other hand, such a design greatly increases the difficulties of producing the required tremendous transfer of heat between gas and air due to the short distance of coaction between air and gas. In other words, the heat transfer surface must be extremely effective.

The heat transfer surface of the rotor may be formed of wire gauze, perforated plates or corrugated metal strips arranged in form of concentric rings or continuous spirals exchanger, 7

A ring shaped 'rotor may be rotated by rollers driven constituting a diskor cylindrically shaped "coil." stiffenfing' elements for the heat transfer surface should then be'ornitted and only 'ring.shaped reinforcements be used.

A'heat transfer surface of this type may be made highly flexible and .capableof adapting. itself by its own weight to the rollers.

fAs one' side oftheheat exchanger operates at very high temperatures; the sealing rolls and the support rolls for the rotor are preferably disposed on the cold side thereof. Means of expansion may be provided between 'th e'hot and cold stationary partsof' the heat exchanger so that tensions and stresses created in these parts by the heat are not transferred directly to the parts of the heat exchanger supporting the rotor.

' According to another embodiment of the invention, a

rotor is employed which is not-only flexible but also of a compressible thickness. In this case pairs ofsealing V rollson opposite sides of the rotor'may be used spaced so tightly that the'rotor portions successively passing between two rolls are compressed relative to the normal thickness of the rotor thereby obtaining a highly efficient sealing between the gas and the air side of the heat by a motor. When, aslsometimes can be the case, the

-rotor is rotated with a' rather high velocity and the gas or the air is" not rotated the rotor would be retarded .by

the traversing air and gas. i 'For instance, when the peripheral velocity of the rotor is IOrneters per second and the flow of gas and air is 50 kg..per second the retarding peripheral'force would be e P= 10=about 60 kg.

wherein g is the gravitational acceleration.

In case it is desirable to use a very light heat transfer surface, for instance in turbines for airplanes which in turn requires a high velocity of rotation, of the heating surface, the aforementioned braking effect can be eliminated by imparting to the air and gas currents a rotation in-the same direction and of the same velocity as that of the rotor. 7 V

With' a sufficient tangential velocity of the flows of air and gas, adriving force instead of a retarding one can be obtained whereby any mechanical driving of the rollers maybe dispensed with. L 7

With a system accordingtoFig-Sfl it will be possible to use pulverized coal as fuel; asash or other solid particles of the fuel will not'reach the turbine.

In order to be able to feed the heat exchanger with combustion gases containing fly ash in spite of the narrow .fiow. canals of the. regenerative mass, the invention fur ther relates to highly etlicient novel particle-removing devices which continuously remove accumulating par air to be heated and one or more outlet ducts 34 for the heated air, Two ducts 33, 34' are shown'in parallel arrangement as can best be seen in Fig. 5. The heat transfer surfaceills has the "general configuration of a dis'ksha'p'ed ring and is flexible, as will be more fully 'explained'hereinafter. Ring 35 is supported by a plurality of conical rollers '37, three rollers being shown. Each roller is journaled near its ends in bearings 38 and 39.

An axle 40 extending from the'inner end of each roller supports a bevel -gear 41*which through free running intermediary gears 36 are in mesh with the bevel gears Fig. 6 shows an arrangement" for collectingand storing collected. Collectors ofgthis type are well known inthe of, adjacent rollers. The rollers: are jointlylrotated i any suitable driverneans such as pulley 42 secured to an axle extending from steamer end of Qdne of-the rollers and coupled with a motor for instance by a belt (r chain drive. As will be notedall the rollers, the/axial lengths of which correspond to the radial width of the 'heat transfer ring, are located within'the cold part of the heat exchanger so that they are'elfectively protected against the'heat. -The bevel' gears are preferably encased" In a housing 43 for protection against dust and heat.

They may further be protected by blocks 44 of'heat insulation material.

inlet pipe 45 and an outlet pipe 46. 7

Forthe purpose of removing solid particles-accumulating on and in the regenerative mass a particle blowing-off device is provided, the construction of which will be more fully described in connection with Figs. l4andl5.

the particles blown off from the heat transfer surface.

This device is shown as comprising a suction funnel 5i) which faces the hot sideof the heat transfer surface 135.

The funnel '50 may have a generally,rectangularicross 'secfionancl'comrnunicates through a pipe 51 with a- .cyclone-type particle collector generally designated by 52.

This particle collector. includes a bin '53 in which the particles removed from the heat, transfer surface'are art. and need not be furtherdescribedin detaiL. Itis of course also possible to equip the. particle collector with:

a suction fan; V r g .The heat exchanger, as. hereinbefore described, operates'as follows. e V The hot gases. are fed to the heat exchanger through 'duct 31 and are discharged cooled from the heat exchanger through duct 32 after passing through the ehlannels of the part of the heat. transferring separating ducts 31 and .32. As a result,.this part of the heat transfer ring is heated'and when it traverses the ducts '33 and 34 its heat is transferred to the cold air entering the. heat exchanger through ductsr33.JThe preheated air is discharged through duct 34. p i

As appears from the previous description,- each part .of the heat surface will alternatingly be he'ated and cooled during its alternating passage through the heat yielding gas and the heat absorbing air. a

Referring now to the design of the heat transferring 35 indetail, Figs; 7' and 8 show a ring composed'jof spiral windings of a'metal band 60, for instance asteel band. The bands 6% are spaced by interposed corrugated bands 6; so as to form a rnultitude of generally. triangular axial channels through the ring. It isof vital importance for the invention that an extremely high number of channels are'formed. To illustrate the actual dimension s of the channels, Fig. 7 shows a fragmentary part of ring In contrast thereto; channels as are 35 in natural size. used for the heat transfer surfaces of conventional regenerative heat exchangers employed in steam installations have approximatelyfthe dimensions of the'channels of Fig. 8. 7

Figs. and 10' show a heat transfer ring in which bands 60 are spaced by bands62 provided with another type 'of corrugations 63 also forming a multitude of separate gaps. 'Fig. 9 again shows approximately the actual dimensions of the channels'formed through the heat transfer surface. a V j 7 'Figs. 1-1 and 12 show an arrangement in which bands 64 are spaced by noses or shoulders 65 to form there- -quired separate gaps. ,Shoulders 65 extend preferably .across the width of bands 64 in the direction of the flow of gas or air. The application of such noses of shoulders .has the advantage that the heat transfer ring can beformed single pspirally wound band. i .The. height of .t e shoulders should not be appreciably greater than the I e V A cooling fluid such as oil may be circulated through the gear housing 43 by means of an in the sealing surfaces.

thickness of the band material. The shoulders may be produced by any suitable means such as rolling the hand.

For reasons which will be more fully. explained hereinafter, it is advantageous to use gaps the radial width of which is small relative to the thickness of the band material. A suitable relation between the width s of the gap and the thickness t or" the band material is:

The windings of which the heat transfer ring is composed should be so loose that individual windings can move relatively to each other to obtain the desired flexibility and to permit the ring to rest with its entire radial width upon the support rollers by its own weight. This has the advantage that the expansion of the ring due to heat cannot cause the ring to warp whereby the inner part of the ring would or might be lifted from the rollers thus making impossible a tight seal between the rotating ring and the stationary parts of the casing.

It will be obvious that the side of the ring formed by the bands which is in contact with the hot gases and the heated air will become appreciably warmer than the opposite side of the ring which is in contact with the cooled gas and the cold air. Such a gradient of temperature within the ring material is desirable as it contributes to the heating of the air to the highest possible temperature and to the cooling of the gas to the lowest possible temperature. In order to attain a maximum gradient of temperature between the hot side and the cold side of ring 35, the individual bands should be as thin as compatible with the required strength, generally, the thickness of the band material should be below 0.5 millimeter. It is further advisable to use a band material having a low heat conductivity such as a nickel-steel alloy. For instance the heat conductivity of steel alloyed with 10-30 percent nickel is reduced by 4640 percent in comparison with that of pure steel. The heat conductivity can also be effectively reduced by alloying the steel with manganese or silicon.

If the ring 35 were rigid the aforementioned gradient of temperature within the ring material would have the effect that the initially approximately flat surfaces of the band spiral 35 would assume a slightly conical or spherical shape. However, such tendency of the ring to deform is rendered harmless by the fact that, as previously mentioned, the individual windings will slide relative to each other by their weight and sink downwardly until they come to rest upon the rollers. One or more pressure and sealing rollers may be provided on both sides of ring 35 as will be more fully described hereinafter. This is suitable when the band spiral is made so light that it is not certain that its weight alone is sufficient to cause it to rest upon the support or sealing rollers.

Figs. 13 and 15 show the arrangement of the support and driving rollers of Figs. and 6 to guide the flexible rotor through the sealing means. These rollers are designated in Figs. 5 and 6 as rollers 37' and are preferably situated on both sides of the air ducts 33, 34 as can best be seen on Fig. 5. The rollers are mounted in a partition wall 70 separating the low pressure cold gas side of the heat exchanger from the high pressure cold air side. For purpose of sealing the rollers each roller is mounted within a casing 71 with the smallest possible clearance. The rollers are further sealed on the cold side of heat transfer ring 35 by fitting packings 72 into the roller casings 71 on both sides of the area of contact of a roller with the heat transfer ring. These packings preferably comprise porous material such as a mixture of gypsum, kieselguhr and magnesia over which ring 35 slides when the rollers are driven. The packings are preferably made to protrude slightly above the scaling surfaces 70" when they are fitted into the recesses provided for this purpose This has the advantage that the heat transfer ring 35 will grind down the packings to .a

10 perfect fit between the ring and the surface of the packin'gs. The packings may be reinforced by a net 73.

Each of the sealing rollers 37 on both sides of the air duct coacts with a heavy substantially rectangular block 74 made for instance of cast iron. This block is disposed within the upper or hot chamber of the heat exchanger opposite to the respective roller 37. The heavy bulk of the block assures that the temperature is approximately uniform within the block and that the same is not deformed by the temperatures to which it is exposed. This block serves as sealing means for the hot side of the rotor. As will be more fully explained hereinafter, the axial thickness of ring 35 is very small so that the heat expansion of the bands of the ring in axial direction is insignificant. As a result and also due to the fact that as previously mentioned ring 35 is highly flexible, body 74 can be positioned with very little clearance close to the ring srrface. face body 74 is preferably provided with saw-tooth shaped grooves 75 longitudinally extending across the direction of the rotation of ring 35. These.

grooves serve further to impede the flow of the high pressure air into the space containing the low pressure gas (see Fig. 15). Instead of the labyrinth sealing formed by the grooves, suitable porous packing material may also be provided as described above. Body 74 is supported by a partition wall 76? separating the hot air side of the heat exchanger from the hot gas side. A packing 76 such as an asbestos packing may be provided between partition wall 759' and body 74.

Instead of sealing by means of the block 74 on the hot side of the heat transfer ring 35, it is also possible to seal by means of sealing rollers built into the casing like the arrangement on the cold side.

Due to the high temperatures to which the rollers on the hot side of the heat exchanger are exposed it is advantageous to cool these rollers. Fig. 18 shows such an arrangement. According to this figure the rollers 95 on the hot side of the heat exchanger are mounted Within a casing forming a coding jacket 96. A suitable cool ing medium is circulated through the channels of the cooling jacket. In order to obtain a uniform temperature of the rollers 95 it is advisable to bring the cooling medium to the boiling point as a boiling cooling medium has practically the same temperature over the entire surface to be cooled in contrast to a cooling medium below'the boiling point and heated during its passage through the cooling jacket. The steam from the cooling medium may be condensed and then recirculated. It is also sometimes advantageous to provide hollow rollers 95 and to circulate the cooling medium through the rollers also.

Each of the rollers 37 and 37 has at both ends a flange 80 and 81 respectively (Figs. 6 and 14) serving as guiding means for the ring.

The outer and inner periphery of the heat transfer ring are encompassed by reinforcement bands 82 and 83 respectively made of stronger material than the band material used to form the channels of the ring. The material used for rings 82 and 83 has prererably a high heat conductivity to maintain a uniform distribution of temperature within said rings and to avoid an excessive deformation by heat.

Figs. 14 and 15 further show a blowing device for removing dust and other solid particles accumulating on and Within the heat transfer ring 35. This blowing device is shown as a channel 85 extending across the radial width of ring 35 and a nozzle 86 also extending across the radial width of ring 35 and facing the same. As can best be seen on Fig. 15, the width of the nozzle is controllable by a nozzle plate 87 detachably secured to a portion of partition wall 76. Opposite the nozzle, that is within the hot gas space of the heat exchanger, a funnel or duct 50 serves to catch dust or other particles loosened by the jet flow of air through nozzle 86. The air under pressure required for the blowing device is supof about one millimeter.

plied'through a duct 88 branched ofiithe ductfor the cold air which is'under high pressure.

To increase the sealing between the inner ends of the rollers 37 or 37' and the gear casing 43, the band 83 encompassing the inner. periphery of ring'35 may be guided by rolls generallydesignated by 90, Figs. 16 and 17, which are preferably circumferentially spaced about the inner periphery of ring 35. Rolls 90 are shown as being composed of an outer metalringi91 and an inner ring 92 made of heat resisting insulation material such as porcelain or a hard ceramic. These rolls are rotatably supported on a rod 93 and serve 'to take up the radial pressure to which the heat transfer ring 35 is subjected, primarily due to the high pressure within the air side of the heat exchanger.

A condition 'for guiding a heat transfer surface between two not yielding sealing means as described above (Figs. .13, 14, 15 and 18-besides its being flexible-is that its extension in the direction of the fiow is very small,

so that the heat expansion of the surface in this direction becomes insignificant.

Furthermore, as has been mentioned, it is desirable that the overall dimensions of the heat transfer surface are as small as possible. This entails that the'heat transfer factor of the heat transfer surfaceisvery high. ,In

the heat exchangers according tomy aforementioned U. S. Patents 2,178,481 and 2,227,836 it is attempted to attain a high heat transfer factor by providing channels through the transfer surface having a hydraulic radius The hydraulic radius is determinedby the equation V circumference.

A hydraulic radius of one millimeter corresponds to a diameter of four millimeters for cylindrical channels or to a width of 42 mm. for a canal of a rectangular crosssection and a relation between its sides from 1:1 tolzOO.

I have now found that channels having a hydraulic radius of about one millimeter make it impossible to obtain the desired high heat transfer factor and, according to the present invention, channels for the passage of air or gas through the heat transfer surface are provided the hydraulic radius of which is only between 0.30 and 0.025 millimeter. The most advantageous hydraulic ,radius of the channels for a flexible heating surface has been found to be between R,,=0.l5 to 0.05 millimeter. I

The reasons which have led to the use of such extremely narrow channels are apparent from the graph of Fig. 19, which shows the heat transfer factor a as a'func- Rex- wherein v is the velocity; d is the diameter'of the tubes,

and 'y the kinematic viscosity. 7

Let it be assumed that the initial diameter d of the tube corresponds'to the point a on the graph and-that- Without changing the flow velocity the diameter of the tube is reduced to. the point 12 on the graph then the heat transfer coefficient decreases to a fraction of the original value. For this reason it is, customa'rypractice .to select a value for the diameter of the tube or in other words the cross-sectional area of, the heat transferring channel whichis greater thanthe value corresponding to point a on the graph, that is Re=2,300. Ho wever, it is'p ossible artificially to produce turbu- "len ce in the gas current. The turbulence which is con- I "12 sideredfor the'flow in a smooth tube and which follows to Reynolds rule, is produced by friction. At Reynolds numbers less than 2,300, friction alone is not sufiicient e to produce'onmaintain a turbulent flow. If, however,

at a Reyonlds number below 2,300, turbulence is pro-, I .duced by special design of'the heating surface,such as is applied in regenerative airpreheaters, the curve of; as a function of the width of the channel takes'on a form according to the point-dash line A.-C. of Fig. 19. But,

an attempt to enforce such a form causes, an extra con sumption of energy (increased resistance) and can; be

carried out only within certain limits with areasonable.

consumption of energy. In the case of a very loW'Reynolds number, the current (flow) remains laminareven when. flowing around sharp edges as the formation of eddies is-deadened by the viscosity forces in the gas. These forces are small with conventional'dimens'iohslof the flow channels but become entirely dominating with very small dimensions. Therefore, with respect to' flow resistance and consumption of power it is not profitable to carry these principles of construction too far,

The present invention solves the hereinbefore outlined problem by refraining from any attempts'tomake the, gas flow artificially turbulent and by selecting a crosssectional area of the flow channel so small that the difficult region of .transition between a turbulent and a. laminar gas-flow is avoided. The invention operates in the region above-the pointDof the graph Fig; 19, that is at the left side of the graph in which very highh eat transfer coefficients can be obtainedin spite of .a laminar flow. While for instance the heat transfer coeflicient increases very little. within a turbulent field in response to a decrease of. the diameter of the tube or flow channel andgeven decreases in the region of transition, that is between points a and b, the heat transfer coeflicient increases strongly in response to a decrease of the tube diameter within the laminar region as soon asthe region This occurs when the so-called Reynold coefficient, etc. sion is halved the heat transfer co'eflicient is doubled. This of'triansition passed. The reason of this phenomenon is that within the laminar region the. heat transfer (20-, eflici'ent is conversely proportional to the diameter of the. tube or to the width of the channel. Consequently, for very smalldimensions of the flow channels a further decreaselof thecross-sctional width of the channel-to one half. willres ult in a doubling of theheatitransfer Hence, whenever the aforesaiddimenincreasein the heat transfer coefficient continues towards an-infinite value until the dimensions of the tubes decreasetowar'ds zero.

The great advantage of this method of increasing the heat transfercoeflicient is that it is not attained at the expense of an increase in the flow resistance and the consumption of energy. The required length in the direction of the flow for a given heat transfer efiiciency and an unchanged velocity of the flow decreases conversely to the square ofthe width of the channel. Consequently,

when Ithe. width. of the channel or tube is reduced .to one-half, the. required length 'of the channel or tube is decreased to one-fourth,;etc. This relation between the width of the channel or tube and the required length thereof is of'vital importance for' the considerations upon which the present invention is based. As previously mentioned, one of the objects of the invention is to extend the heat transfersurface as little as possible in thedirection of the flow as a small extension in the s aid direction permits efiiciently to solve the also afore- .mentioned very diflicult sealing problem. Practical tests and calculations have shown that the extension of the heat transfer surface in the direction of the flow should 7 be and isless than 5% of the outer diameter ofthe ring shaped heat transfer-surface, the ring 35 in the previously described exemplifications of the invention. The sub 'sequent'fconcrete examples will demonstate how such .ashort extension can be obtained for a heat efliciency of Let it be first assumed that with a heat transfer surface according to any of the Figs. 9 to 12 the width of the gap is 0.3 millimeter and that the effect of the spacers or shoulders 63 and 65 may be disregarded. The hydraulic radius of the gap is then 0.15 millimeter. To facilitate the calculations, the complicated heat transfer conditions near the inlet edge of the gap are also disregarded. In

the deeper regions of the gap both the velocity-gradient of the flow and the gradient of temperature are substan tially stabilized and the heat transfer coefiicient can be calculated with sufiicient accuracy in a simple and reliable manner from the heat conduction coefiicient of the air and the gas respectively. As the flow at the places mentioned is fully laminar, the Reynolds number being far below the critical value, the heat transfer is effected exclusively by conduction. The heat conduction coefficient of the gas and the air within the heat exchanger are calculated as the mean value of the gas and air side at the above mentioned temperature; it is 0.04 kcal. per meter, hour and C. These figures refer to the amount of heat which would be transferred by heat conduction exclusively, between a surface giving off heat and a surface taking up heat separated by a cube of immovable air having a side of one meter and when the diiference of temperature between the said surfaces is 1 C. Such a transfer of heat is of course very small if compared with the heat transfer which would actually take place as the result of a convection flow/ thermal circulation in a cube of the aforesaid size. The heat transfer due to thermal circulation is about 100 times higher and may be forced to become still many times higher as the result of increased turbulence. It is therefore understandable that a turbulent flow has hitherto been considered as the only feasible way of obtaining an acceptable heat transfer coetficient. To employ a laminar flow for the purpose of heat transfer in connection with gaseous media having such a low heat conduction coefficient as air and gas has been considered as being impractical. The present invention abandons entirely the principles previously generally accepted and provides by the adoption of extremely small and truly critical dimensions of the flow channels a heat exchanger which employs a laminar flow and which is nevertheless far superior to heat exchangers as hitherto known and employing a turbulent flow.

In the previously given example the distance for the heat transfer within the cube of air was one meter. When now the distance between the surface giving off heat and the surface taking up heat is reduced to one centimeter, 100 times as large an amount of heat is transferred or 100 0.04:4 kcal. per meter hour and C. That is, still a low value in comparison with the heat transfer coefficient obtainable by a turbulent flow and forced draft.

As mentioned before, the heat transfer surface of the heat exchanger according to the example now under discussion has a radial width of the gap of 0.3 millimeter. The sheet of air which flows through the heat transfer surface is heated on both sides. The greatest distance the heat must be transferred within the mass of air is from the heating surface to the center of the gap, that is a maximum of 0.5 s=0.15 millimeter. However, the mean distance for the transfer of the heat within the gas is considerably smaller. As a detailed calculation of the mean distance is too involved for the purpose of this discussion, it will be hereinafter simply assumed to be half the distance from the heating surfaces to the center of the channels that is 0.25 s. The heat transfer coefiicient can now be calculated from the equation a 0.25 s s =533 kcal/m h C.

s being calculated in meters. In the inlet portion of the gap, the distance for the heat transfer is considerably less than 0.25 s due to the fact that no larger portion of the air sheet has been heated as yet. Only an extremely thin layer near the heating surface has been heated so that the heat transfer through this layer to the portion of the air still to be heated within the gap occurs very quickly. Hence, the actual transfer of heat within the inlet part of the gap is considerably higher than has been calculated.

The fraction of the total axial length of the gap that may be considered as the inlet part of the gap within the meaning of the previous discussion is not a fixed value as it depends upon the velocity of the flow. The higher the velocity is the deeper will penetrate the layer of air unaffected by the temperature of the heat transfer surface Consequently, the portion of the heat transfer surface which aifords a higher heating transfer coefficient than previously calculated is correspondingly increased. When the air flow reaches the region of the gap within which the gradient of temperature is substantially stabilized, the heat transfer coefiicient becomes independent of the flow velocity and dependent entirely upon the width of the gap and the heat conducting coefficient of the medium. In other words, the previously calculated heat transfer coefficient represents a minimum of transfer which in actual practice will be considerably exceeded. Because of the safety margin obtained in this manner, the resistance to heat conduction in the plate and other reducing factors may be disregarded in the subsequent calculations. The heat conduction coefiicient of steel previously referred to as hand material for the heat transfer ring 35 is several hundred times that of air. Consequently, the resistance of the band material to heat conduction is negligible in comparison to the resistance to heat conduction within the layer of air or gas, particularly according to the principle of regenerative heat transfer owing to which the heat conduction in the plate does not take place from one side to the other but to and fro through the same surface layer. A considerably more reducing factor, in view of the high heat efiiciency and the small band width here under consideration, is the backfiow of heat from the warm edge to the cold edge of the band. in view of this phenomenon it is advantageous to use an alloyed steel having a low heat conducting factor as already mentioned, A material having a heat conducting coefficient higher than one-half of that of pure iron is not suitable as material for a heat transfer surface according to the present invention. In actual practice an alloy may be selected having a heat conducting coefficient of one-fourth or less of that of pure iron. Another reducing factor is the change of the temperature of the plate during a rotation of the rotor.

Disregarding this factor too for the sake of simplicity the total heat transfer coefficient between gas and air may then be calculated by the following equation:

k= =267 kcaL/m h C.

i l irL a l. 533 533 To compensate for the surface portions of the heat transfer surface facing the sealing means which are substantially ineffective for the transfer of heat, 7% of the ring 35 is 0.6 millimeter.

calculated value is preferably added. Accordingly, a required single-sided surface of the band of 1000 m? will.

the gap is selected which corresponds to an air inlet velocity of 10 m. per second at the face of ring 35 and "when the gas inlet velocity is selected twice as high, the

required face'area on the air side of ring 35 is 2.2 m.

[at three atmospheres and onthe gas side of :the ring 3.5 m? at one atmosphere. By adding 7% of the respective areas to compensate for the surface portion of ring .35. covered by the sealingmeans a total face area of 6.1 m. for the ring;35 is obtained. This corresponds to a total length of the band of 1.667 times 6.l=l0.l69

meters. The required width of the band will then be which equals approximately $5 meter. 7 The fiow'resistance is inthe magnitude of 600 millimeters of water column on the air side and 750 millimeters on the gas side of the heat exchanger. 7

When the aforesaid figures are compared with conventional designs employing a heat transfer surface composed of plates or strips having a thickness of 2 millimeters and forming gaps having a width of 10 millimeters,

similar calculations show that the heat' transfer surface must have an axial extension of 2 meters to obtain the same heat transfer surface per m? of face surface. Furthermore, due to the lower heat transfer coeflicient,

.with a gap of 10 millimeters width the actual total heat transfer surface would have to be many'times larger so p that the required dimensions would be excessive in view 'Of the. sealing problem.

Figs. 20, 21 and 22 show an arrangement .in which in contrast to the previously described exemplifications of the invention the flexible heat transfer ring 35 is not rotated by the power driven rollersbut by the velocity of the air or gas current. The ring 35' is supported by rollers 100 on the cold side of the ring. The gas or air is directed against ring 35 by a plurality of circumferentially spaced blades 101. These blades are so slanted that the tangential component of movement'imparted to the current will produce the desired rotation or ring 35.

The arrangement of Figs. 20, 21 and 22 should be visualized as constituting part'of a heat exchanger as has been described in connection with the previous'figures. Fig. 23' shows an arrangement similar to Fig; 22 and in addition the figure shows counter-rollers 102 on the hot side of ring 35 in the air duct. Y r As is evident from the previous description, due t the multitude of extremely small channels through heat transfer ring 35 great care must be taken-that these small channels are not clogged by the accumulation of ash, dust or precipitation and also to prevent that solid particles reach the turbine to any harmful extent. It has been found that a blowing-off device as described'in connection with Figs. 13 and is not always s'uflicient to remove precipitations from the heat transfer ring to a satisfactory extent. Figs. 24 to'27 inclusive describe a cleaning device which may be used instead or in addition to the blowing-oif devices of Figs. 13 and 15 and which is of an extremely high efiiciency. V 7

Before describing the cleaning device of Figs. 24 to 27 in detail, the phenomenon which is employed in the device shall be explained in connection with Figs. 28 to 31.

Figs. '29 and 28 show a U-shaped tube 105 the right hand shank of which is higher than the left hand shank by the height'h. Let it be assumed that both shanks are initially tilled with powdermaterial and tha't the density. of.

the' material is approximately equal in the two: shanks. When new a flow of gaseous fluid is slowlyradmitted through a pipe 106 into theU-shaped tube at the' midpoint of the bight thereof aquantity of powderparticles will be forced out from the open end of the 'left hand'shank of the tube, while thematerial in theright'hand shank will sink dueto the fact that theinitialtotal weight of the left hand column is somewhat less than the total initial. "weight of .the right hand column. Once this 'unequal process of discharge has been initiated it willrapidly become more and more pronounced as the density and hence the weight of the left hand column is reduced in compari son to the right hand column. Consequently, the right height h as indicated in Fig. 29. r

' Figs-30 and 31- show a similar arrangement in which the left hand shankof a tubeforms a'funn'el-107 and-the hand column will drop until finally it .will stop'at-the.

' right hand shank is extended and bentat 108 to face the funnel. Initially, the'left'handshank is completely filled and the right hand shank substantially empty." When now gaseous fluid is slowly admitted through pipe 106' the powder material within the left hand shank will gradually sink down into the right handshankdue to'the uneven distribution of weight and a circulation is started as indicated in Fig. 31. The'circulatory movement of the particles can be facilitated and accelerated by making the diameter D of the right hand tube larger than the diameter d of the left hand tube.

Reverting'now to Figs. 24 to 27, there is shown a riser pipe'110 communicating through a bend 111 with apipe 112 leading from a collecting receptacle 113 disposed below a portion of the rotating'heat transfer ring '35@ Riser pipe 110 leads to a distributor generally designated by 114 situated above ring 35 and opposite to receptacle .113. 'As can bestbe seen'in Figs. 25 and 26, the distributorcomprises a casing 115 of rather narrow substantially-rectangular shape which extends radially across the ring 35 and which communicates with pipe 110.v .There are provided within casing 115 a plurality of guide'baffles 116 to distribute the current of powder circulating within the device across the casing 115. The casingcommuni- 'cates with a narrow duct 117 also having a cross-sectional length equal to the radial width of ring'35 and 'continued in a curved duct 118 formed within a guide member 119 disposed closely adjacent to the hot side of the heat transfer ring 35; As can best be seen on Fig. .27, duct 118 is tangentially disposed relative to the face surface of ring 35. 'Tangentially opposite" to' duct 11'8 anwo'ppositely against'which particles are directed and louvered'on the opposite side at 121;

Air or gas under pressureis fed from a suitable source of supply through a pipe 122 which communicates with the midpoint of the bendllL Fi'g. 24.' .A second pressure pipe 123 may be provided which communicates withthe 'pipe 110slightly above the bend 111. .This'se'cond' pipe serves as a starter pipe to facilitate the initial movement 1 device. However, the heaviest possible particle material of the material within the device as has been described in connection with Figs. 28.to 31. i f I A certain quantity of suitable discrete particles such as finely grained clean sand may be initiallyffilled inithe is preferred such as a metal powder. It is advantageous to use a metal alloy the melting point of which is below the temperature in the'turbinein order to protect itin the event that particles are carried by the rotor to the air side of the preheater. The above described device should, of course, be placed at the gas. side.

The operation of the cleaning device, as just described,

is, as follows:

To start the device, air under pressure is forced into the device initially through both pipes 122 and 123. Ma

result, the discrete particles within" the system will begin to rise through pipe and willfcirculate through'ducts' 

